Fluid bearings, especially gas bearings, are known in the art and which are effective for the purposes intended. For example my prior U.S. Pat. Nos. 3,249,390 and 3,476,451, whose disclosures are incorporated herein by reference, describe fluid bearing systems and gas bearing systems which operated satisfactorily but which were difficult and relatively expensive to manufacture and which had a relatively large profile and were relatively heavy due to the large profile.
Those structures included spaced bearings and a rotor or shaft in which the forward bearing was relatively fixed, during operation, with respect to axial and radial and angular motion. In one form, there was no spring support for one bearing while the other was supported at spaced points along the outer periphery. In another form, the hydrodynamic form, the forward support bearing was fixed axially. For the purposes of this invention, the terms shaft and rotor refer to the rotating part while the term bearing refers to the fixed, i.e., non-rotating element which supports the shaft for rotation on a fluid or gas film. While the rear bearing of these prior structures permitted limited axial movement, and limited radial and angular movement, the assembly as a whole was limited in movement in that only one end of the rotor or shaft was capable of limited movement. As a result, in the event of a significant force axially of the rotor or shaft, or angularly or radially of the moveable end of the rotor, the forward end of the rotor would tend to contact the forward supporting bearing and bottom out on that bearing. The result was a catastrophic failure of the bearing assembly because of the failure to mount the supporting bearing in a resilient manner, to be described.
Thus, there are limitations in that prior art structure as described in my prior patents. More specifically, the support bearings were supported at three separate points disposed 120 degrees to each other such that the forward support bearing was essentially fixedly supported in operation, but with relatively severely limited movement. The forward support, during operation, was urged axially of the rotor to seat against the opposed face of the supporting housing by annular shoulders on the support. In order to move radially, it was necessary for the forward bearing to overcome the friction of the shoulders against the housing. Seals between the support bearing and the housing were in the form of piston rings, which were split, somewhat similar to those of automotive piston rings. Due to the three point support and the fact that the support bearing was seated against the housing, the forward support bearing was not capable of angular movement with respect to the rotor axis and to the extent that there was any movement at all, there was relatively high friction.
As a result the friction, coulomb friction, of the forward support bearing assembly, the friction of the total bearing assembly was relatively high and was used as a dampening effect. It was believed then that the coulomb friction, i.e., the friction which has to be overcome to effect relative movement of non-lubricated parts, was operative to provide a dampening effect between the forward support bearing and the remaining structure to overcome the radial oscillatory motion of the assembly.
The total friction of the assembly, however, was relatively high, although it was thought at that time to be acceptable. The system was intended to be a two degree freedom of movement system including two springs and two masses. The rotor was one mass and the support bearings were the second mass while one spring was the gas film and the other spring was the radially located separate three spring system, as described, between the support bearing and the housing. Since one of the support bearings was fixed axially in operation and capable of some radial movement at a relatively high friction, the effectiveness of the system as a two degree of freedom system was somewhat less than desired. This became apparent when one attempted to use the prior described bearing at relatively high rotational speeds and where axial loads were placed on the rotor.
In such a case, the entire structure must be capable of responding relatively rapidly to changes in load and pressure in order to operate satisfactorily. To achieve this, it was determined that axial, angular and radial motion was desirable in order to provide the relatively rapid response needed to accommodate changes in load and pressure, and especially in response to thrust loads. Further, such motions had to be achieved with relatively low friction, i.e., relatively low coulomb friction, so that the parts responded quickly and relatively smoothly rather than being relatively non-moveable until the relatively high coloumb friction was overcome.
In addition to the above, it is known in the prior art to use O-rings as a seal between the bearings and the support housing in which the bearings are mounted. The difficulty with O-rings is that they are usually compressed radially to function as a seal and are of limited resiliency in a radial direction and tend to perform erratically as the temperature increases. Further, O-rings tend to degrade over time due to chemical changes in the material of which the rings are made. Normally, these O-rings are not considered to be resilient support elements, but more of a sealing element with the result that they tend to be relatively high friction elements, even if coated with low friction material such as polytetrafluoroethylene or other fluorocarbon or low coefficient of friction materials. As such, the inherent design of the O-ring is somewhat controlled in the sense that one has to design the environment in which they are used to match the structure to the character of the O-ring. This imposes some severe limitations on the use of O-rings, especially if it is necessary to provide a resilient support operative over a wide range of operating temperatures, for relatively long periods of time, and which is capable of adjusting to different types of relative movement between the mating parts.
In general, gas lubricated bearings are normally assumed to have a given spring constant and a small amount of inherent damping. The spring constant may be measured by plotting a force displacement curve while the damping coefficient may be determined by ascertaining the decrease in amplitude as a function of time of a bearing system after it is subjected to an impact load.
However, gas bearings are different from more conventional bearings in the sense that there is a finite time lag between the initial application of a displacement force and the time required for the bearing to reach a steady state condition. Upon extensive investigation of experimental evidence, it was noted that this time constant or time lag is difficult to measure because it is obscured by the bearing's damping. This time constant can be calculated and becomes larger if pocket type bearings are used. It can also be shown by computer simulation that, in general, the longer the time constant (first-order lag), the more unstable the system becomes. This time lag is caused by the fact that a finite amount of time is required for the gas to flow out of the bearing or into the bearing until a new pressure distribution has been reached.
It is known that a pocket type bearing has a greater restoring force, i.e., the force tending to establish a balancing equilibrium condition, an since a resiliently mounted fluid bearing is more stable, a bearing system employing pockets may be utilized. Thus, a resilient mounted system may be designed for a gas bearing in which the added advantage of the greater restoring force associated with pocket type bearings can be used even though they have a longer time constant.
As pointed out in the patents previously referred to, there are applications in which the use of conventional lubricants in a bearing presents significant problems. For, example, in cryogenic applications, ordinary petro-lubricants and the like, or silicone lubricants become unusable due to the fact that they become viscous or solid. In oxidizing environments, some lubricants are totally unacceptable due to possible catastrophic failure. Materials such as graphite may be used, but are of limited utility. All of these, and other factors are discussed in the prior patents referred to previously.
In instances in which the gas bearing is used to support high speed spindles, in the range of 15,000 and preferably 80,000 to 130,000 or higher RPM, such as those used in circuit board drilling machines or other high RPM accurate drilling equipment, the debris caused by such operations produces "cuttings" which may be quite abrasive or otherwise hostile. Properly designed in accordance with this invention, rotational speeds as high as 450,000 RPM may be achieved. The cuttings usually are relatively light weight and are in the form of relatively fine dust which may infiltrate the bearing or the spindle and which may shorten the life thereof because of the introduction of relatively abrasive particles into the relatively small space between normally precision and manufactured parts which rotate relative to one another. For example, roller bearing spindles suffer the problem of a relatively high wear due to the presence of fine abrasive particles unless some special and sometimes relatively heavy and expensive structures are provided to prevent entry of such abrasive particles into the working space between the relatively stationary and rotating elements of the bearing and spindle assembly. Such debris may cause serious damage to O-ring seal systems.
In the case of that type of environment, the use of a gas lubricated bearing, in accordance with the present invention offers singular advantages because the bearing structure is relatively immune from such relatively hostile environments since the exhaust of the gas used in the bearing prevents most of the debris from entering the bearing or spindle housing. Nonetheless, there are other concerns regarding the use of gas bearings due to some of the problems traditionally associated therewith, some of which have already been discussed.
Gas lubricated bearings and spindle assemblies are known in the art and some of these have attempted to overcome the problems noted. However, these structures, some of which are described in patents noted and some of which are currently commercial products, such as the gas bearing available from Federal Mogul Corporation under the trademark WESTWIND, have not solved the problem of a light weight and reliable structure capable of operating at comparatively high rotational speeds in environments which generate hostile and abrasive products. For example, the Westwind ball bearing system weighs about five or more pounds, while the Westwind gas bearing system is heavier and requires air filters for the incoming air and air driers to reduce moisture thereof as well as refrigeration systems to cool the incoming air.
One of the advantages of the present invention is the provision of a resiliently mounted gas lubricated bearing which is essentially immune to such a hostile environment since the exhaust gas discharged by the bearing prevents essentially all of the debris from entering the bearing or spindle housing. The obvious advantage is that there is no need for relatively heavy sealed and grease lubricated assemblies, while providing a relatively light weight structure capable of operation at comparatively high speeds and in environments in which corrosive products tend to adversely affect the performance of even sealed lubricated assemblies.
More recently, efforts have been made to overcome the above noted problems through the use of conventionally lubricated bearings. The results have been products of relatively low useful life and which are relatively heavy and bulky in order to provide the encapsulation shielding needed to reduce the effects of such corrosive and harmful products. Even so, the results have been marginal at best as compared to this invention.
In the case of gas lubricated bearings in general, lubrication is achieved by containing a thin film (about 0.0005 of an inch) of gas between an accurately machined shaft journal and the bearing. The result has been to provide bearings for special applications which are substantially insensitive to super extremes in temperature. Further, since there is no contact between the relatively rotating parts, there is no coulomb friction, heating or rubbing, and consequently no wear from such effects. As is understood, properly designed gas lubricated bearings may be operated at very high rotational speeds, with long life and stable performance characteristics. Nonetheless, the prior designs have yet to reach that level of performance in terms of light weight and high rotational speeds which gas lubricated bearings are capable of achieving.
Gas lubricated bearings may be of the hydrostatic or hydrodynamic type. Both types are contemplated by the present invention. In a hydrostatic bearing, gas is continuously supplied to the bearing interspace under a predetermined pressure. In hydrodynamic gas bearings, the gas lubricating film is self-maintaining when relatively high tangential velocities are reached, relative to the bearing surface and shaft and may, if desired, be totally isolated from other sources of gas. This capability to be sealed off causes hydrodynamic bearings to be attractive in applications such as gyros and the like where the maintenance or use of a source of pressurized gas may not be feasible.
Yet, the machining tolerances for bearings of the hydrodynamic type are significantly more stringent than for the hydrostatic type of bearing. In the hydrodynamic bearing, the bearing interspace gap is usually less than about 0.0001 of an inch. This relatively small gap gives rise to detectable viscous friction at high bearing speeds and is manifested as drag on the rotary shaft. Further, the self-maintaining of the film pressure inherently limits the versatility of the load capability and the selection of critical (resonant) angular speeds of the bearing.
In the case of hydrostatic bearings the machining tolerances are relaxed somewhat as compared to hydrodynamic bearings but are nonetheless stringent. Moreover, the gas feed into the bearing interspace must, in some cases, be angularly symmetric in order to support the rotary shaft thus to preclude imbalance with respect thereto. Even so, the hydrostatic approach provides a more versatile and stable bearing because of its control ability and larger spacings.
The prior art devices in the gas bearing area, including the patents previously referred to, achieve support for axial and radial thrusts or loads by the use of a pair of juxtaposed spaced cylindrical surfaces and a pair of juxtaposed spaced radially arranged surfaces wherein one surface of each pair is on the shaft and its juxtaposed counterpart is or is supported on a stationary frame or housing. There is no appreciable cooperation between the two bearing parts, their supporting forces being mutually orthogonal.
A generic deficiency of the prior art gas bearing structures, including those of the patents identified, is that regardless of the machining tolerances and regardless of the care with which the shaft is loaded in use, a finite rotary imbalance exists which causes an oscillation in conjunction with the elastic restoring force of the supporting gas film. While these bearing assemblies are thought of as being two degree of freedom systems composed of two springs and two masses in that the rotor and bearing are the masses and the gas film and supposedly resilient mounting are the spring system, they do not behave as two degree of freedom systems.
In practice, the supporting gas film has a relatively low "spring" constant and the resonant critical speed of the shaft is so low as to be a severe limitation on high rotary speed applications of gas lubricated bearings. The severity of the resonance problem is caused by the fact that the near zero viscous friction of the gas film affords near zero damping of the oscillating bearing. Consequently, the bearing structure tends to oscillate without limit until the moveable element strikes the stationary bearing or bushing. Typically the resulting coulomb friction precludes driving the shaft above the critical frequency or causes destructive wear or both. Where the coulomb friction is relatively high, as is the case in the structures of the patents referred to, the response of the bearing system as a whole to changes in load, pressure and the like is relatively slow and this tends to promote bearing "crashes" or limits the use of the bearing in terms of rotational speed. In effect the system is essentially fixed and relatively incapable of axial, radial and angular movement to compensate effectively for such variations during operation of the bearing system.
Another difficulty with the gas bearings of the prior art is loss of gas film pressure with the result that there is metal-to-metal contact at certain rotational speeds. It is believed that this is due to "whirl", a phenomenon which is caused by the rotating shaft being displaced, due to its weight, off-center with respect to the axis of the stationary bushing. In effect, the shaft rotates on an axis such that one end of the shaft circumscribes a circular track, with the result that the shaft is closer to the bushing at one point than at others, and the shaft experiences angularly out-of-balance viscous drag. This drag, or the reaction which it produces, is usually in the of the shaft rotation and effectively causes a rotation of the shaft which is effectively angularly unsymmetrical. The whirl rotation, being due to the reverse-directed drag, is usually the same as the shaft rotation and is seen by the gas support film as a reduction in shaft velocity. Thus, with increases in the whirl velocity, since it is in the same direction of shaft rotation, a significant reduction in the support capability of the fluid film may occur with the result that there is metal-to-metal contact between the rotating and non-rotating parts at high speeds. While the restoring force and the whirl resistance of the film may be influenced, to some degree, by the fluid pressure in the case of a hydrostatic bearing, and the mass of the shaft may be minimized in order to increase the critical speed, these represent comprises in the stability and load capability of the bearing and do not represent a generally applicable solution to the problem.
In general, the solution to the problem, as set forth by this invention, is to make the interior of the bearing structure resiliently self-adjusting so that it responds rapidly to these events. This may be accomplished by relative angular, rotational and axial movement of the support bearings. By contrast, the structure of the patents previously identified is such that the support bearings are not moveable angularly at the front end and the bearing has a relatively high friction in a radial direction, with little if any movement in an angular direction. While that system appears to be a true two degree of degree of freedom system, in structure and operation it is not.
Thus, it is an object of the present invention to provide an improved fluid, preferably a gas, resiliently mounted lubricated bearing and method which are not subject to the disadvantages of the prior art systems referred to above.
Another object of the present invention is to provide a bearing structure of the type to be described in which the bearing assembly which provides the fluid or gas film for rotation of the shaft is resiliently mounted in essentially a free floating condition.
An important object of this invention is the provision of a fluid and preferably a gas bearing system, capable of operation at relatively high rotational velocities, in which the bearing assembly is flexibly and resiliently mounted to a support structure so that the relative movement of the bearing and the shaft is such that the gap between the fluid supported shaft and the bearing is maintained to permit the entire structure to compensate for relative axial movement of the shaft and to compensate for relative angular and radial movement of the shaft with respect to the assembly which supports the bearings and the shaft.
It is another object of this invention to provide a gas bearing system in which a pair of juxtaposed surfaces, resiliently mounted, provides resilient and compensating support for both axial and radial thrust of the rotatable shaft.
Still another object of this invention is the provision of a bearing system of the type described in which the gas or other fluid film thickness in both a radial and axial direction is adjustable by the resilient axial or radial relative movement of one of the relative supporting surfaces through a unique resilient mounting thereof.
Another object of this invention is the provision of a fluid and preferably gas bearing system in which frictional damping is coupled to the radially oscillating shaft without solid-to-solid contact between the rotating shaft and the supporting structure by the provision of a unique flexible and resilient mounting with permits relatively rapid adjustment of the relative parts to maintain the required fluid and gas gap between the relatively moving parts.
Yet another object of this invention is the provision of a bearing system of the type described in which the allowable amplitude of radial oscillation of the shaft at critical frequency, without solid-to-solid contact, may be increased without increasing the quiescent thickness of the fluid supporting film through the use of a unique resilient supporting system.
Still another object of this invention is the provision of an improved three-degree of freedom resiliently mounted bearing system of the type described in which the thickness of the fluid, preferably a gas, film is essentially self-adjusting by the provision of an effectively self-adjusting film thickness achieved through the use of a unique resilient mounting system of the bearing system which allows the weight and profile of the bearing system to be reduced while achieving relatively high rotational speeds through the achievement of essentially the required compensation due to relative axial, angular and radial movement of the relatively moving parts.